1. Field of the Invention
The present invention relates generally to lubrication of rotary seals in order to prolong seal life and reduce running torque, wear and self-generated heat. The rotary seals are suitable for environmental exclusion and lubricant retention, and more particularly relate to lubrication of seals that may be exposed to high differential pressure, to a high percentage of compression, and to rotation in the clockwise and counter-clockwise directions.
One aspect of the preferred embodiment of the present invention includes a hydrodynamically lubricating seal lip geometry suitable for both compression-type (i.e., interference-type) and flexing, cantilever lip-type rotary seals.
2. Description of the Related Art
Oilfield mud motor sealed bearing assemblies, rotary steerable tools, rotary blowout preventers and other oilfield equipment present a number of difficult rotary sealing challenges, as detailed in commonly assigned U.S. Pat. Nos. 4,610,319; 5,195,754; 5,230,520; 5,738,358; 5,823,541; 5,873,576; 6,007,105; 6,036,192; 6,109,618; 6,120,036; 6,227,547; 6,315,302; 6,334,619; 6,382,634; 6,494,462; and 6,561,520. For examples of mud motor sealed bearing assemblies, see U.S. Pat. Nos. 3,730,284; 4,372,400; 4,476,944; and 6,439,866. For examples of rotary blowout preventers, see U.S. Pat. Nos. 4,281,724; 5,178,215; 5,224,557; 5,647,444; 5,662,181; 6,016,880; 6,354,385; and 6,554,016.
The current state of the art seals for such equipment are the family of hydrodynamically lubricated seals manufactured and sold under the above-referenced commonly assigned patents. These seals—known in the industry by the registered trademark “KALSI SEALS”—typically have a sinusoidal wave-type hydrodynamic geometry for introducing a lubricant film into the dynamic sealing interface between a dynamic seal surface and a rotatable member. Although such seals perform well under certain conditions, there are other conditions in which their performance is deficient. One such set of difficult conditions is where the seal is subject to both clockwise and counter-clockwise rotation and where increased lubricant film thickness is desirable. An example of such an application is the mud motor. Mud motor rotary seals are subject to clockwise rotation during normal drilling and counterclockwise rotation due to counter-clockwise wind-milling that can occur as the drillstring is lowered.
In the prior art seals having a sinusoidal wave-type hydrodynamic geometry, the film thickness is greatest in the region of the sinusoidal waves because much of the lubricant that enters at the wave leading edge simply leaks out at the wave trailing edge. As a result, the dynamic sealing interface is less well lubricated toward the environment-side which causes undesirable wear, torque and seal-generated heat—particularly when one or more of the following conditions are encountered:                Thin viscosity lubricants, which are less than optimum for lubrication;        High differential pressure across the seal;        High ambient geothermal temperature;        Shaft materials and configurations that interfere with heat transfer;        High initial seal compression;        High rotary speeds; and        Closely spaced redundant seals that concentrate seal-generated heat.        
Running torque is related to lubricant shearing action and asperity contact in the dynamic sealing interface. Although the prior art hydrodynamic seals run much cooler than non-hydrodynamic seals, torque-related heat generation is still a critical consideration. The prior art seals are typically made from elastomers, which are subject to accelerated degradation at elevated temperature. For example, media resistance problems, gas permeation problems, swelling, compression set and pressure-related extrusion damage all become worse at higher temperatures.
The prior art seals cannot be used in some high speed or high-pressure applications simply because the useful temperature range of the seal material is exceeded due to seal-generated heat. Additionally, the bi-directional rotation prior art seals do not provide sufficient lubrication if the sealing lip incorporates harder materials—such as reinforced polytetrafluoroethylene (“PTFE”) based plastic.
Typically, seal life is ultimately limited by susceptibility of the seal to compression set (i.e., permanent deformation), abrasive wear, and extrusion damage. Many applications would benefit from a hydrodynamic seal having the ability to operate with greater initial compression to enable the seal to tolerate greater mechanical misalignment, run-out, tolerances, compression set, and wear. Many applications would benefit from a seal that is better lubricated in extreme operating conditions, and suffers less wear. Many applications would also benefit from a cooler-running seal that would sustain less temperature-related loss of modulus of elasticity, and would therefore have increased high pressure extrusion resistance.
FIGS. 10-13D of U.S. Pat. No. 6,109,618 show a convex hydrodynamic inlet geometry—a rounded surface shape that in a circumferentially aligned cross-section would have a radiused appearance that is similar to the leading edge of a traditional sled runner. Seals manufactured in accordance with FIGS. 10-13D of U.S. Pat. No. 6,109,618 have an inherent but subtle limitation which is this: a given size of a substantially circumferentially oriented convex hydrodynamic inlet cannot provide the same rate of convergence with respect to the shaft across a family of seals for different shaft diameters.
Shaft diameter has a significant effect on the rate of convergence of a given convex hydrodynamic inlet size. Smaller diameter shafts converge with the inlet at a faster rate, creating more abrupt convergence. This affects lubricant wedging efficiency. With the same convex inlet size, a smaller diameter seal based on FIGS. 10-13D of U.S. Pat. No. 6,109,618 will lubricate less well and produce less flushing action, and a larger diameter seal will lubricate better and produce more flushing action.
FIG. 17 of this specification schematically shows the uncompressed relationship between various shaft diameters and circumferentially oriented hydrodynamic inlet radii. Even though FIG. 17 is not a representation of installed convergence, it provides insight into the relationship between shaft diameter and inlet convergence.
As can be readily seen from FIG. 17, the rate of convergence of a 2.00″ convex inlet radius against a 2.75″ diameter shaft is approximately double what it would be against a 16.50″ shaft. As can also be seen, the rate of convergence of a 2.00″ convex inlet radius against a 0.50″ diameter shaft is even more abrupt. As a result, better lubrication and a higher flushing rate occur with a 16.50″ shaft, compared to when the same inlet size is used on a 0.50″ or 2.75″ shaft.
As can be seen from the dashed line on FIG. 17A, if the rate of convergence of a 2.00″ inlet radius against a 16.50″ diameter shaft were to be duplicated on a 0.50″ diameter shaft, it is now understood (as a result of recent research by the inventors performed under United States Department of Energy contract no. DE-FG02-05ER84206) that the inlet radius would have to be concave instead of convex. In other words, it is now understood that small diameter seals have to incorporate concave inlets in order to achieve the same degree of hydrodynamic lubrication as larger diameter seals that have convex inlets. U.S. Pat. No. 6,109,618 does not contemplate the use of a concave inlet. In fact, the manufacturing methods used to create the tooling for the seals represented by FIGS. 10-13D of U.S. Pat. No. 6,109,618 cannot accommodate a concave inlet.
Miniaturization Problems
As downhole tools are miniaturized, a number of design issues arise because some parameters cannot be scaled down linearly due to practical manufacturing constraints, tolerances and application specific requirements.
As the rotary seal is scaled down, the smaller radial cross-sectional depth results in less dimensional compression for a given percentage of initial compression. Tolerances cannot be scaled down below certain practical limits, thus contributing to larger compression variability. Shaft stiffness also scales down non-linearly, contributing to larger radial deflection and runout. This causes higher compression on one side of the seal, and lower compression on the other side.
A certain minimum level of dimensional compression is required so that a seal can accommodate tolerances, misalignment, seal abrasion, and compression set without losing sealing contact with the shaft. The factors stated above dictate that the initial percentage of compression of miniature seals be increased, compared to larger cross-section seals, to achieve the necessary dimensional compression.
Contact pressure at the dynamic sealing interface is related to the percentage of compression and the modulus of elasticity of the seal material. A higher percentage of compression leads to higher seal-to-shaft interfacial contact pressure, which reduces hydrodynamic lubricant film thickness, causing higher friction, seal-generated heat and wear. The compression limitations of the prior art seal designs impose corresponding limits on the minimum cross-section that is practical to manufacture. To compensate for this, seals capable of stronger hydrodynamic action in either direction of rotation are needed.